Loop-scavenged two-stroke internal combustion engines

ABSTRACT

A loop-scavenged two-stroke internal combustion engines with an intake valve ( 7 ) engaging a seat ( 10 ) for fresh air intake, and an exhaust valve ( 8 ) engaging a seat ( 13 ) for combustion gas exhaust, is disclosed. The valves are arranged in such a way that the fresh air intake scavenges a substantial part of the burnt gases. In at least one of the valves, the valve surface ( 21 ) located downstream from the valve face ( 9 ) and the surface ( 23 ) of the downstream extension of the seat ( 10 ) are configured in such a way that they form a substantially isentropic diffuser.

FIELD OF THE INVENTION

The present invention concerns an improvement to two-stage internalcombustion engines with loop scavenging of the type

having at least one variable volume working chamber delimited by acylindrical wall in which a piston slides, the mobile top face of thepiston and a fixed cylinder head,

operating in accordance with the two-stroke cycle, with a loopscavenging system via the cylinder head, controlled by at least oneinlet valve cooperating with a seat, preferably of generally conicalshape, to cause the working chamber to communicate cyclically with aninlet cavity communicating with means for supplying air to the engineand by at least one exhaust valve cooperating with a seat, preferably ofgenerally conical shape, to cause the working chamber to communicatecyclically with an exhaust cavity communicating with the combustion gasexhaust system of the engine,

and in which said inlet and exhaust valves are disposed so that the airentering the working chamber through the inlet valve causes scavengingof at least a substantial part of the burned gases in the chamber andtheir evacuation via the exhaust valve.

BACKGROUND OF THE INVENTION

In engines of the above type the difference AP in the gas pressurebetween the means supplying the engine with air at pressure P and theengine combustion gas exhaust system is relatively low and in practiceis imposed by the specifications of the supercharged air supply means.

Scavenging can take place only during a limited part of each cycle,another and large part of the cycle being devoted to compression andexpansion of the gases renewed in the chamber.

As a result the geometry and the operation of the inlet and exhaustvalves play a decisive role in the efficiency, the power and the speedof the engine.

Increasing the size of the valves rapidly runs up against a geometricallimit imposed by the dimensions of the cylinder head while increasingthe valve lift, that is to say the distance the valve moves away fromits seat, and the lift speed, which are determined by the profile of thecams opening and closing the valves, is rapidly limited by mechanicalconstraints imposed by the permissible contact pressure between the noseof the cam and the components that it actuates.

This limits performance, i.e. the permeability of the cylinder head, theefficiency of use of the air passing through the cylinder head, i.e. theratio between the mass of air enclosed in the working chamber at the endof scavenging to the mass of air passing through the cylinder head, andthe scavenging efficiency, i.e. the ratio between the mass of air andthe total mass of gas enclosed in said chamber at the end of scavenging.

Patent application EP-A-0 673 470 (or U.S. Pat. No. 5,555,859 or WO95/08052) has made it possible to improve significantly on the abovelimitations by providing a single inlet valve and a single exhaustvalve,

said inlet and exhaust valves being coaxial circular cylinders,preferably coaxial with the cylindrical wall of the working chamber, thecoaxial arrangement being such that the inlet valve is outside theexhaust valve,

the seat of the inlet valve being attached to the cylinder head andoriented so that the pressure of the drive fluid contained in theworking chamber exerts a force that tends to press said valve onto itsseat, said seat being in the immediate vicinity of the periphery of theupper part of said cylindrical wall inside which the piston slides, andin contact with the cylinder head,

said exhaust valve having a tubular part the inside wall of which slideson a fixed hub carried by the cylinder head, to which it is sealed bysealing means, and the end of which towards the chamber has a bearingsurface coaxial with said tubular part so that it can cooperate with aseat provided inside the lower part at the same end of the chamber assaid inlet valve, enabling communication of said exhaust cavity with theworking chamber, by virtue of the annular space delimited radially bythe inside wall of the inlet valve and by the outside wall of theexhaust valve.

This arrangement optimises and controls scavenging and doubles theactual lift of the exhaust valve because the inlet and exhaust valveslift in opposite directions.

If means are provided to cause the inlet air passing through the inletvalve to rotate, axi-symmetrical centrifugal layering can be achievedand the fuel can be injected into a hot central area that is relativelyimpoverished in oxygen, to obtain the advantages described in the abovepatent and in patent application FR-A-2 690 951.

BRIEF SUMMARY OF THE INVENTION

The present invention proposes to improve further the performance ofengines having concentric inlet and exhaust valves, in particular asdefined above, enabling a choice between reducing the scavenging timeand consequently increasing the usable expansion stroke of the engineand therefore its efficiency, or, for a given angular duration ofscavenging during the upward travel of the piston, a high permeabilityof the cylinder head enabling the rotation speed of the engine to beincreased.

Another objective of the invention is to reduce significantly wear ofthe valves and their seat and in particular the inlet valve and itsseat, the exhaust valve and its seat being generally better protected bythe effects of lubrication by the carbon deposits caused by combustiongases moving towards the exhaust.

The invention consists in a two-stroke internal combustion engine withloop scavenging, preferably of the compression ignition type, of thekind described in the preamble, and preferably in which the inlet andexhaust valves are coaxial circular cylinders, preferably coaxial withsaid cylindrical wall, preferably with the inlet valve outside theexhaust valve, and preferably having the other features of the valves ofconcentric valve engines of the type described hereinabove,characterised in that, for at least one of said inlet and exhaust valvesthe surface of the valve downstream of the bearing surface of saidvalve, in the direction of flow through it, on the one hand, and thesurface of a part extending the seat of said valve, with which saidbearing surface cooperates, and also situated downstream of said seat,on the other hand, are configured to constitute a substantiallyisentropic diffuser discharging into the cavity downstream of the valve.

“Isentropic diffuser” means a divergent nozzle in which the flow of gasthrough the nozzle is slowed and compressed virtually isentropically.

The outlet into said downstream cavity has a discharge section greaterthan the geometrical flow section of the valve in the fully openposition between the bearing surface of said valve and its seat.

“Geometrical flow section of the valve” means the minimum unrestrictedflow section between the lifted valve and its seat. This section remainsin the vicinity of the bearing surface, i.e. the areas of contactbetween the closed valve and its seat, but its axial position can varywith the lift of the valve.

Accordingly, flow downstream of the valve is effected in a diffuserwhose section increases progressively, at least on approaching thecavity downstream of the valve, i.e. the working chamber in the case ofan inlet valve and the exhaust cavity in the case of the exhaust valve,the discharge section of the diffuser, i.e. the section through whichthe diffuser opens into said cavity, being greater than the flow sectionof the valve in the fully open position between the bearing surface ofthe valve and its seat.

In a preferred embodiment, the ratio between the discharge section wherethe flow from the valve enters the cavity downstream of the latter, inthe direction of flow, and said geometrical flow section of the valve inthe fully open position is at least equal to the critical ratiocalculated for the value of the ratio of the pressures of the fluidflowing in said valve on either side of the latter during normaloperation of the engine.

The critical ratio is defined as that for which the speed of the flowreaches the speed of sound at the throat of the fluid flow in thevicinity of the valve bearing surface; it can easily be calculated usingthe equations of isentropic diffusion.

If the valve forming the diffuser is a tubular inlet valve opening intoa cylindrical chamber, the surface profile of the valve downstream ofits bearing surface is preferably configured so that it progressivelybecomes substantially parallel to the direction of the cylindrical wallof the chamber in which the piston slides.

The bearing surface of the inlet valve, in conjunction with the seat,can then advantageously define an angled path progressively widening inthe meridian plane and with advantage initially oriented outwards, i.e.towards the wall of the chamber, so as progressively to become parallelto the wall of the chamber.

In the case of a tubular exhaust valve the profile of a part of thevalve downstream of its bearing surface is preferably configured at itsoutlet so as to be substantially parallel to the interior cylindricalpart of the inlet valve that forms the seat of the exhaust valve.

In a preferred embodiment deflector means are provided in the air supplypassage that delivers air to the tubular inlet valve to impart to theflow of air a rotary component so as to send through the flow area ofthe valve and then the part forming the diffuser a substantiallyisentropic flow of air with a rotary movement procuring a centrifugaleffect tending to hold the air against the wall of the cylindricalchamber in which the piston slides to obtain the advantages described inpatent application FR-A-2 690 951.

The access passages to the inlet valve are preferably inclined to thegeometrical axis of the cylinder at the exit from said inlet cavity andtowards the piston to reduce deflection of the flow in a meridian planein order to minimise head losses.

Deflector means such as fixed deflector blades can be disposed either inthese passages or even on the inlet valve itself, if necessary. Byincreasing the effective permeability of the cylinder head by virtue ofthe isentropic diffusion of the flow, the invention limits thedisadvantageous head loss that inevitably results from the rotationimparted to the air by the deflector means, which are generally inclinedat an angle near 45°.

The invention also consists in an engine as defined in the preamblepreferably including a tubular inlet valve having a bearing surface thatis preferably generally conical in shape cooperating with a valve seatcarried by the cylinder head, the bearing surface of the valve bearingon the seat along a circular line in a plane perpendicular to the axisof translation of the valve, characterised in that the valve and theseat downstream of said circular line of bearing engagement between thebearing surface and the seat as defined when the pressure in the chamberis low or nil are adapted so that on the occasion of cyclic deformationof the valve by forces due to the pressure of the gases the diameter ofthe circular line of contact decreases so that the bearing surface ofthe valve pivots about its bearing engagement with its seat and rollswithout sliding on the latter.

The deformation of the valve due to the action of the pressure of thegases can then be exploited with advantage to prevent sliding leading towear of the bearing surface of the valve by achieving an effect ofrolling against the surface of the seat, by virtue of the deformation ofthe valve, when the pressure increases as a result of compression andthen combustion.

This rolling contact without sliding is possible only because of thehollow structure of the valve on each side of the line of contact.

To achieve this result the conjugate surface at the bearing surface andpreferably also the surface at the seat advantageously have profileshaving a point of inflection, the line of contact, i.e. of bearingengagement, moving in the vicinity of this point of inflection when thepressure varies.

With surfaces of the above kind, for example S-shape surfaces, when thevalve is not loaded by the pressure, or only slightly loaded, the lineof contact is below the point of inflection. When the valve is highlyloaded, when the gases are at the maximal pressure, for example, it ison or slightly above the point of inflection.

The conjugate surfaces are preferably adapted to form, downstream of thebearing surface and the seat in the direction of flow of the fluid, adiffuser procuring substantially isentropic flow as defined hereinabove.

BRIEF DESCRIPTION OF THE DRAWINGS

Other advantages and features of the invention will become apparent onreading the following description given by way of non-limiting exampleand referring to the appended drawings, in which:

FIG. 1 represents a schematic view in axial section of a prior artengine shown with the inlet and exhaust valves closed and the piston attop dead centre while combustion is taking place in the working chamber;

FIG. 2 represents a schematic view in axial section identical to FIG. 1but with the valves in the maximally open state and the piston nearbottom dead centre while the working chamber is being scavenged;

FIGS. 3 and 4 represent a schematic view of the geometrical shapes ofinlet and exhaust valves with their respective seat designed inaccordance with the invention to enable flow through them to diffuseisentropically downstream of their respective throats, i.e. downstreamof the maximal constriction of the passage available to said flow. InFIG. 3 the inlet valve is open and the exhaust valve is slightly open.In FIG. 4 both valves are maximally open;

FIG. 5 represents a schematic view of the lower part of the exhaustvalve with a floating ring disposed between its inside surface and thehub;

FIG. 6 represents a schematic view of the inlet valve actuating means;

FIG. 7 represents a schematic sectional view of the inlet valveincluding a space containing a heat-conducting fluid; and

FIG. 8 represents a schematic view of the conjugate profiles of theinlet valve and its seat in one preferred embodiment of the invention.

DETAILED DESCRIPTION OF THE INVENTION

Refer first to FIGS. 1 and 2.

The prior art engine described in application EP-A-0 673 470 is atwo-stroke diesel engine with loop scavenging comprising a cylinder 1 inwhich slides a piston 2 and which is closed at the top by a cylinderhead 5.

The cylinder, the piston and the cylinder head delimit a variable volumeworking chamber 3 in which combustion takes place when the piston isnear top dead centre, as shown in FIG. 1.

The cylinder head 5 has a circular cylindrical fixed central hub 6attached to it the axis 23 of which is preferably coincident with thatof the cylinder and the piston and inside which there is a fuelinjector, not shown, on the axis of said hub and discharging along theaxis of the combustion chamber 4 forming part of the working chamber 3.

The engine also includes a generally tubular inlet valve 7 and agenerally tubular exhaust valve 8, said valves being concentric alongthe common axis 23 and the exhaust valve being inside the inlet valve.The inlet valve 7 has at the bottom a bearing surface 9 cooperating witha seat 10 formed in the lower part of the cylinder head 5. Concentricwith and outside this bearing surface, an inlet cavity 11 distributesair to the valve from a conventional air supply device (not shown). Thecavity 11 advantageously communicates with the inlet valve 7 viapassages 12 oriented to impart to the inlet air flow a rotationcomponent about the common geometrical axis 23 of the various componentsof the engine.

These passages can be delimited by two coaxial conical surfaces in thecylinder head, fixed deflector blades being disposed as close aspossible to the exit into the working chamber.

The tubular inlet valve 7 has a circular cylindrical inside surface,preferably with a conical surface in its lower part coaxial with theaxis 23 of the valve and forming a seat 13 that cooperates with thebearing surface 14 of the tubular exhaust valve 8.

The tubular inlet valve 7 also has a cylindrical tubular body 15 whichslides in a bore in the cylinder head 5. The tubular exhaust valve 8also has a cylindrical tubular body which slides on the fixed centralhub 6. Oil passages are formed between the inside cylindrical surface ofthe body of the exhaust valve 8 and the outside cylindrical surface ofthe fixed central hub 6 to enable cooling and lubrication of the facingcomponents.

Because of their different diameters, the two coaxial circular tubularvalves delimit an annular passage 16 through which the exhaust gases areconducted from the working chamber 3 to the exhaust cavity 17 whichcommunicates with the exhaust system, not shown, of the engine.

The two valves 7 and 8 are operated hydraulically, as described inapplication EP-A-0 673 470, by virtue of cyclic variations in thepressure of a hydraulic liquid enclosed in two cavities 50 or 60 ofconstant volume but having variable surface areas delimited by a drivepiston 51 or 61 cooperating with a camshaft 53 coupled to the main shaftof the engine and by a receiving piston 52 or 62 respectively attachedto the inlet valve and the exhaust valve, and which includes adequatereturn spring means.

The receiving pistons 52 and 62 are disposed so that the inlet andexhaust valves are actuated in opposite directions, the tubular inletvalve 7 opening downwards, i.e. towards the piston, and the tubularexhaust valve opening upwards.

With the bearing surface 14 of the exhaust valve cooperating with itsseat 13 on the inside surface of the inlet valve, and with the twovalves moving in opposite directions, the opening of the exhaust valveis clearly increased by the lift of the inlet valve.

An engine of the above type operates in the following manner:

When the piston 2 is propelled towards bottom dead centre by the gasesin the working chamber after combustion of the fuel, and therefore atthe end of expansion of the working chamber, the exhaust valve is openedto enable the pressure in the working chamber to fall below the pressurein the inlet cavity 11 (“exhaust puff”) to prevent the gases flowingback towards the inlet circuit (“counter-scavenging”). The inlet valveis then opened to scavenge the working chamber, which consists insubstituting air for the combustion gases.

The inlet air enters the flow section of the inlet valve delimitedbetween the bearing surface 9 and the seat 10, having had rotationimparted to it previously in said passages 12. The air therefore entersthe working chamber in the space delimited by the lateral wall of thecylinder and the lower part of the valve. Because of the rotation of theair about the axis 23, the air streams entering the working chamber areinclined to said axis 23, forming a layer of air along the cylindricalwall moving towards the piston and rotating about this axis.

At the same time the hot gases which are concentrated near the axis ofthe chamber 3 and of the combustion chamber 4 escape via the flowsection between the seat 13 and the bearing surface 14 of the exhaustvalve. During the first phase of the upward movement of the piston 2 thecombustion gases are therefore largely evacuated and replaced by air. Inthe second part of the upward movement of the piston 2 the valves areclosed and all of the gases contained in the chamber 3 are thenprogressively compressed to the state of maximum compression in thecombustion chamber 4 into which fuel is then injected under pressure,which ignites the fuel and starts a new engine cycle.

The times at which the valves close can advantageously be adjusted toobtain inside the available volume above the piston 2 a mass of air thatis rotating and therefore centrifuged towards the periphery andsurrounding a smaller mass of hot combustion gas in the central part,from the previous cycle and retained in the chamber during scavenging,with the result that injection takes place into this central part, whichwill procure the advantages described in patent application FR-A-2 690951.

In accordance with the invention, and as shown in FIGS. 3 and 4, thelower part of the tubular inlet valve 7 downstream of the bearingsurface 9 cooperating with the conical seat 10 (in the direction of flowof the air through it) is extended downwards, i.e. towards the piston,by a skirt 21 having symmetry of revolution and coaxial with said valve,the outside surface of which, i.e. the surface at the greatest distancefrom its axis 23, is a circular cylindrical surface the meridian profileof which merges at the upstream end tangentially with the surface of thebearing surface 9. At the downstream end it preferably terminatesparallel to the axis 23.

In the same manner the conical seat 10 is extended regularly by acircular cylindrical surface 22 around the axis 23 the meridian profileof which is tangential—at the upstream end—to the conical seat andparallel—at the downstream end—to the axis 23.

The facing circular surfaces 21 and 22 therefore delimit an annularpassage having symmetry of revolution and the discharge section of whichincreases regularly from the minimal flow section Sc (with the valve inthe maximally open position) called the throat of the valve and in linewith the seat 10 and the bearing surface 9 to the maximal flow sectionSd at the bottom of the skirt 21.

In accordance with the invention, the meridian profiles of the facingsurfaces 21 and 22 are designed to constitute an “isentropic diffuser”when the valve is in the fully open position. By “isentropic diffuser”is meant a passage whose flow section, increasing in the direction offlow, is such that the flow through it, having been previouslyaccelerated and expanded until it passes the throat of the valve, isthere decelerated and recompressed quasi-isentropically [i.e. with nothermal losses at the wall and with conservation of the total pressure(cut-off pressure) all along the flow], to the static pressuredownstream of said valve.

The progressive increase in the flow section in the diffuser must beneither too small—because friction at the walls would then becomeexcessive, leading to a drop in the total pressure of the flow—nor toolarge—because the flow then separates from the wall, also leading to adrop in the total pressure. For a conical diffuser, for example, theoptimal angle characterising the progressive increase in the flowsection is known to be around 7° to the axis of said cone.

An arrangement of the above kind has the following advantages:

The flow, which is deflected relative to the axis of the cylinder onpassing through the throat of the valve, is regularly straightened so asto be directed parallel to the axis of the cylinder towards the piston(with a tangential component, if momentum is imparted to this flow onpassing said valve).

For a given pressure different AP between the inlet cavity and thecylinder, the value of which depends on the efficiency characteristicsof the turbocharger, and a given temperature T and a given pressure P inthe inlet cavity 11, the flowrate Q through the valve is increased inthe ratio S_(d)/S_(c) of the maximal discharge section S_(d) at the exitfrom the diffuser to the minimal section S_(c) at the throat relative tothe flowrate Q* that would pass through the same valve without itsdiffuser.

The increase in the flowrate is nevertheless limited by the fact that,expanding on passing through the throat of the valve, the flow isaccelerated and is then recompressed and decelerates in the diffuser.However, for a given section at the throat (determined by the geometryand the maximal lift of the valve) there is a maximal value of theoutlet section (S_(d))_(m) of the diffuser for which the speed of theflow at the throat of the valve reaches the speed of sound. If the exitsection of the diffuser were greater than this critical section(S_(d))_(m) the flow would separate from the wall beyond said criticalsection and there would no longer be an isentropic diffusion of the flowbeyond the critical section and consequently no increase in the flowratethrough the valve.

The critical section of course depends on the expansion ratio ω=P(P−ΔP)between the inlet cavity and the cylinder. The theoretical expressionfor this quantity is:

[S_(c)/S_(d)]_(critical)={2/(γ−1)·[(γ+1)/2]^((γ+1)/(γ−1))·ω^(n)·(ω^(n)−1)}^(½)/ω

with: n=(γ−1)/γ and γ=C_(p)/C_(v)

C_(p) and C_(v) being the specific heats at constant pressure and atconstant volume of the gaseous fluid concerned.

For example: γ=1.404:

ΔP/P [S_(d)/S_(c)]_(critical) 0.05 2.228 0.10 1.622 0.15 1.366 0.201.222

This means that the flow through the valve of the invention can beincreased 62% relative to a conventional valve if the pressuredifference on either side of the valve is 10% of the total upstreampressure. On the other hand, there is no point in increasing the exitsection of the diffuser beyond this critical ratio, as the flow wouldthen be stuck at the speed of sound on passing through the throat of thevalve.

The diffuser is said to be matched to the throat of the valve if thespeed of sound is reached at said throat and if the flow diffusesreversibly, i.e. without separating from the wall, as far as the exitsection.

Thus if the diffuser is matched to the valve in the fully open position(with a discharge section increased 62% relative to the section at thethroat if the pressure difference across the valve is 10% of thepressure upstream of the valve) it would clearly not be matched forsmaller lifts of the valve. If the valve is lifted halfway, for example,the flow will separate from the wall of the diffuser in the section ofthe diffuser whose flow section is equal to half its exit section.However, the facing profiles 21 and 22 can be organised so that thediffusion is as near perfect as possible even during lifting of thevalve. This considerably increases the velocity of the flow at thethroat and this considerably increases the effective permeability of theinlet valve for a pressure difference between the inlet pressure due tothe supercharging means and the pressure in the exhaust, which remainsconstant. The flow of air entering via the inlet valve is thereforeconsiderably increased.

It should also be pointed out that the diffuser is extremely efficient,including at the instant the inlet valve begins to move away from itsseat and the flow area at the level of the seat is still very small. Theefficiency, i.e. the increase in permeability, is immediately obtained,even at low engine operating speeds and when starting the engine, inother words at time when, in the conventional arrangement, scavenging ismost difficult.

Likewise, in the exhaust valve of the invention, as represented in FIGS.3 and 4, the circular cylindrical surfaces constituting the inside wall25 of the inlet valve and the outside wall 24 of the exhaust valve,situated downstream of the bearing surface 14 of the exhaust valve andthe seat 13 of said valve formed on the inside wall of the tubular inletvalve, delimit an annular passage 26. The meridian profiles of thesefacing surfaces 24 and 25 are designed so that the annular passage 26constitutes a divergent passage from the minimal section Sc at thethroat of the valve to the maximal value Sd where it joins the exhaustpassage 17 and so that this divergent passage is an isentropic diffuserin the sense defined above.

Similarly, the ratio between the exit section Sd of the diffuser and thesection Sc at the throat of the valve in the fully open position ispreferably at most equal to the critical ratio calculated for thenominal value of the relative pressure difference across the valve.

The position of the throat of the valve, defined by the minimalgeometrical flow section of the annular passage, can vary relative tothe position of the bearing surface 14 and the seat 13 of said valve andin accordance with the degree to which the valve is open.

In FIG. 3, for example, in which the exhaust valve is shown slightlyopen, it can be seen that the throat of the valve, with the minimal flowsection Sc, is in the immediate vicinity of the bearing surface 14 ofthe valve and its seat 13.

In FIG. 4, on the other hand, which shows the two valves in the fullyopen position (and with the flow section of the valve increased becauseof the downward movement of its seat 13 on the inlet valve), it can beseen that the position of the throat of the valve of minimal section Scis well above its bearing surface, in the rectilinear part of theannular passage, which is highly favourable to obtaining good diffusionof the flow (it is well known that it is particularly difficult toobtain perfect diffusion in a curved passage).

The skilled person can determine conjugate profiles 21 and 22 for theinlet valve and 24 and 25 for the exhaust valve either by calculation orby experiment. The profiles are preferably designed so that thedivergent passage following on from the throat of the valve constitutesas near perfect as possible an isentropic diffuser, in the sense definedabove, when the valve concerned is in the fully open position. Todetermine the ideal profile it is necessary to take into account thefact that the flow has an axial component (“discharge velocity”) and atangential component imparted to it on passing the deflector means suchas blade 6.

Refer now to FIG. 5.

The inlet valve has a large diameter and centres naturally, when closed,on its seat on the cylinder head. The smaller diameter exhaust valvebears on the conical seat 13 on the inlet valve and can therefore beoff-centre by a non-negligible amount relative to the central hub 6 onwhich it slides. In the case of large bore engines, given the tolerancesfor the manufacture and stacking of the parts, this eccentricity can beas much as several tenths of a millimeter.

Under these conditions, to assure a good seal between the insidecylindrical surface of the exhaust valve 8 and the outside surface ofthe central hub 6, a floating ring 28 can advantageously be fitted thatis able to move laterally relative to the hub. The floating ring 28 canbe accommodated in such a fashion as to be able to move laterally with asmall clearance in a groove formed between a shoulder 29 on the hub 6and a counter-shoulder 36 on a part 31 which is also part of the hub 6,the outside cylindrical surface of the ring 28 providing a sliding trackfor a sliding seal or packing 32 in the lower inside part of the exhaustvalve 8. The sliding track obviously has an outside cylindrical surfaceof sufficient height to enable sliding of the packing 32 throughout thelifting of the exhaust valve.

Refer now to FIG. 6.

Any known valve actuation means can be used to actuate the valves, forexample the inlet valve 7, such as mechanical actuation by a camshaft,for example, or electromagnetic actuation synchronised with rotation ofthe main shaft of the engine. In any event, hydraulic actuation meanscan advantageously be used, as in patent application EP-A-0 673 470,which consist in a deformable cavity of constant volume filled with ahydraulic liquid such as the lubricating oil of the engine, for example,and having a first chamber 34 of variable volume delimited by a cylinderhead and in which slides an actuator piston 35 cooperating with acamshaft 36 coupled to the main shaft of the engine and whichcommunicates via a passage 37 with a second chamber 38 of variablevolume delimited by the bore in which the cylindrical outside surface ofthe inlet valve 7 slides, which has a shoulder 39 serving as a receiverpiston so that when the nose of the cam 36 actuates the drive piston 35the oil, deemed to be an incompressible liquid, expelled through thepassage 37 from the first chamber 34 into the second chamber 38, causesthe receiver piston 39 to descend and thereby the valve 7 to be opened.Return movement in the upward direction can be effected by return means,for example a spring or preferably pneumatic return means consisting ofthe compression of the air contained in a cavity 40 one face of which isalso delimited by a shoulder 41 of the valve 7 acting as a return pistonsurface. If the volume formed by the cavities 34 and 38 and the passage37 is too full of oil, for example after an oil leak into the cavity orthermal expansion of the oil, the valve will not drop back onto itsseat. If there is insufficient oil, for example through leakage to theexterior, contact between the cam 36 and the roller of the piston 35will be interrupted, which will cause impacts in the actuating means.

The invention avoids these drawbacks by means of an automatic device fortaking up clearance providing a small diameter passage 42 that candischarge into the cavity 34 and is connected to the low-pressure oilsupply means 43, the outlet from the narrow passage 42 into the chamberbeing disposed so that it is cyclically covered and uncovered by themovement of the piston 35, its location being such that, when the piston35 with its roller is released to return to its initial position afteractuation by the cam 36, the outlet is uncovered and places the cavityfilled with oil in communication with said low-pressure oil supply means43, whereas it is very quickly covered when the piston 35 begins to bemoved downwards by the cam to start lifting the valve.

Refer now to FIG. 7.

In the engine in accordance with the invention the inlet valve 7 has alarge area exposed to the combustion gases with the result that it isimportant to cool the valve effectively. Likewise, because of the highperformance that can be obtained with an engine of the above kind, theexhaust valve can advantageously be rigorously cooled.

In accordance with the invention, the valve can advantageously be madewith an elongate annular cavity inside it substantially exposing theshape of the valve and descending to a point near the free end 27 of thevalve 7 so as to extend largely inside the cylindrical tubular part 15of the valve. This cavity is partly filled with a fluid 44 that is agood conductor of heat, for example sodium which is in the liquid statewhen the valve has reached its operating temperature. In this way heatcan be evacuated outwards into an area in which it is easy to cool thevalve. Moreover, the large surface area swept by the air duringscavenging enables transfer of heat from the inside surface of the inletvalve during combustion to the outside surface of said valve duringcompression. This transfer can be obtained either by conduction or byconvection in the heat-conducting fluid.

A similar cavity can be provided in the exhaust valve in order to conveyheat from the lower end of the valve to the water or oil cooling meansof the valve.

Note that this technique, using a heat-conducting fluid such as sodiumpartly filling a cavity within the thickness of the valve and more orless espousing its outside surface and consisting in extracting heat inthe hot part of the head of the valve in order to transfer it to thestem of the valve where cooling means are disposed, is known in itselfbut of very low efficiency. With a valve of conventional shape there isa disproportion between the surface area that receives the heat (the“tulip” or valve head) and the surface area where the heat can beevacuated (the valve stem).

On the other hand, with a tubular valve as used in the invention, theseproportions are reversed and there is a very large tubular surface areafor evacuating heat extracted from the head of the valve by means of anappropriate cooling system.

Refer now to FIG. 8.

In a tubular inlet valve, such as a valve 7, for example, associatedwith the tubular exhaust valve 8 of which it carries the seat 13, theforces due to the pressure of the gases, especially when the piston isnear its top dead centre, are exerted on the inside face of the inletvalve, mainly facing the chamber 4. The valve part above its bearingengagement with the seat 10 will be subjected to tensile stresses andthe free end of the valve below this bearing engagement will be subjectto compression stresses. The resultant of these forces is exerted on thevalve 7 between where it bears on the fixed seat 10 and where it bearson the mobile bearing surface 14 of the exhaust valve 8. The combinationof this resultant of forces due to the action of the gases and bearingengagement reaction forces (represented by the solid line arrows) exertsa tilting torque (the forces of which are represented by chain-dottedarrows) on the inlet valve 7 which therefore pivots about its fixedbearing point, i.e. the seat 10, anticlockwise in FIG. 8.

The conjugate profiles of the bearing surface 9 and the seat 10 of thevalve 7 can be calculated allowing for the radial compression strengthof the part of the valve under the bearing surface 9 and the radialtensile strength of the part of the valve between the bearing surfaces 9and 13, so that the bearing surface 9 of the valve 7 bears on its seat10 at a circular contact line the plane of which is perpendicular to theaxis of said valve, and which can roll without sliding of the seat 10when the pressure of the combustion gases cyclically deforms the valve7. This effect of rolling without sliding can be obtained by imparting arounded profile to the surface of the valve at its bearing surface 9 andto the surface of the cylindrical head at the seat 10, whether the valveis a tubular valve of the type defined in the present invention, that isto say one in which the surfaces downstream of the seat form aquasi-isentropic diffuser, or a conventional tubular valve.

In the tubular inlet valve of the invention, which includes aquasi-isentropic diffuser downstream of its seat, the fact that thelatter is hollow and has an S-shape profile on opposite sides of itsbearing surface 9 lends itself particularly well to obtaining thisrolling without sliding of the circular line of contact of the bearingsurface 9 on the seat 10, which line will migrate upwards (i.e. towardsthe cylinder head) in a plane perpendicular to the axis of said valvebecause of the cyclic deformation of the valve by the pressure of thegases in the combustion chamber. The conjugate profiles of the bearingsurface 9 and the seat 10 of the valve can be determined by experiment,seeking to minimise friction and therefore wear of the parts in contact,or by calculation, using the evolution of the thickness and the shape ofthe valve (and therefore its inertia) as a function of the axialposition of the section plane and of the stiffness of the material fromwhich it is made.

What is claimed is:
 1. Two-stroke internal combustion engine with loopscavenging comprising: at least one variable volume working chamberdelimited by a cylindrical wall in which a piston slides, a mobile topface of the piston and a fixed cylinder heat, said engine operating inaccordance with a two-stroke cycle, with a loop scavenging system viathe cylinder head, controlled (a) by at least one inlet valvecooperating with a seat to cause the working chamber to communicatecyclically with an inlet cavity communicating with means for supplyingair to the engine and (b) by at least one exhaust valve cooperating witha seat to cause the working chamber to communicate cyclically with anexhaust cavity communicating with a combustion gas exhaust system of theengine, wherein said inlet valve and said exhaust valve are disposed sothat the air entering the working chamber through the inlet valve causesscavenging of at least a substantial part of burned gases in the chamberand the evacuation thereof via the exhaust valve, wherein, at least oneof said inlet and exhaust valves and (a) a surface of the at least onevalve downstream of a bearing surface of said at least one valve, in thedirection of flow through the at least one valve (b) a surface of a partextending the seat of said at least one valve, with which said bearingsurface cooperates, and also situated downstream of said seat, and (c)are both said surfaces configured when the at least one valve is in acompletely opened position so that a fluid flow section increasesprogressively in the direction of flow to constitute a substantiallyisentropic divergent diffuser discharging into the cavity downstream ofthe at least one valve.
 2. An engine according to claim 1 wherein adischarge section at a point of entry into said cavity is greater than ageometrical minimal flow section of the at least one valve in a fullyopen position.
 3. An engine according to claim 1: wherein there is asingle inlet valve and a single exhaust valve, wherein said inlet andexhaust valves are coaxial concentric circular cylinders with a commonaxis and with the inlet valve outside the exhaust valve, wherein theseat of the inlet valve is attached to the cylinder head and oriented sothat a pressure of a drive fluid contained in the working chamber exertsa force that tends to press said inlet valve onto its seat, said seat ofthe inlet valve is in an immediate vicinity of a periphery of an upperpart of said cylindrical wall in which the piston slides and in contactwith the cylinder head, and wherein said exhaust valve has a tubularpart with an inside wall (a) which slides on a fixed hub carried by thecylinder head and (b) which is sealed to said fixed hub by sealingmeans, and wherein an end of said exhaust valve towards the chamber hasa bearing surface coaxial with said tubular part so as to be able tocooperate with said seat of the exhaust valve, said seat of the exhaustvalve being formed inside an end facing towards the chamber of saidinlet valve, enabling the working chamber to communicate with an exhaustcavity by virtue of an annular space delimited radially by an insidewall of the inlet valve and by an outside wall of the exhaust valve. 4.An engine according to claim 3 further including means for impartingrotation to inlet air passing through the inlet valve.
 5. An engineaccording to claim 1 wherein a ratio between a discharge section where aflow from the at least one valve enters the cavity downstream thereof,in the direction of flow, and a geometrical minimal flow section of theat least one valve in a fully open position, is at least equal to acritical ratio calculated for a valve of a ratio of pressures of a fluidflowing in said at least one valve on either side thereof during normaloperation of the engine.
 6. An engine according to claim 1 wherein ameridian profile of the surface of the inlet valve downstream of thebearing surface thereof is configured so as progressively to becomesubstantially parallel to a direction of the cylindrical wall of thechamber in which the piston slides.
 7. An engine according to claim 1wherein meridian profiles of an outside surface of the exhaust valve andof an inside surface of the inlet valve downstream of a minimal flowsection of said inlet valve are configured at an outlet of an annularpassage so as to be substantially parallel to an axis common to saidvalves.
 8. An engine according to claim 1 wherein said inlet cavitycommunicates with said working chamber via passages inclined to alongitudinal axis and towards both said chamber and said piston so as toreduce a change of direction of an inlet flow of air in a meridian planeof the engine.
 9. An engine according to claim 8 wherein said passageswhich are inclined relative to the axis comprise a passage between twocoaxial conical surfaces in the cylinder head in which are disposed, asclose as possible to an outlet from said passage, fixed deflector bladesadapted to impart to the flow through said passage a rotation componentaround the axis.
 10. An engine according to claim 3 wherein an insidecylindrical surface of the exhaust valve cooperates with a ring forminga track on which slides a seal packing disposed between the central huband said exhaust valve so as to be able to move laterally relative tosaid hub and assume a position coaxial with the seat of the exhaustvalve disposed in the inlet valve in an event of eccentricity of thefixed hub relative to said seat.
 11. An engine according to claim 1wherein the inlet valve has a shoulder serving as a piston sliding in acylinder and delimiting a variable volume chamber communicating with acylindrical variable volume chamber in which slides a piston cooperatingwith a camshaft in order to raise the inlet valve and wherein saidchamber delimited by said piston is connected to a low-pressure oilsupply means by a narrow passage having an outlet which is cyclicallycovered and uncovered by a movement of said piston cooperating with thecamshaft so that when said piston is released to return to an initialposition thereof after actuation by the cam the outlet is uncovered andplaces the cavity filled with oil in communication with the low-pressuresupply means whereas said outlet is very quickly covered when the pistonbegins to be moved by the cam to begin to lift the inlet valve.
 12. Anengine according to claim 1 wherein at least one of the valves is atubular valve having an elongate internal cavity espousing the shape ofthe at least one valve and with the internal cavity partially filledwith a heat-conducting fluid enabling evacuation of heat to a tubularupper part of the at least one valve.
 13. An engine according to claim 1wherein a bearing surface of the inlet valve and the seat of the inletvalve downstream of a circular line of contact therebetween when thepressure in the chamber is low or nil, are both adapted so that oncyclic deformation of said inlet valve by forces due to a pressure ofgases a diameter of the circular line of contact decreases so that thebearing surface of the inlet valve pivots about a line of bearingengagement with the seat of the inlet valve and rolls without slidingthereon.
 14. An engine according to claim 13 wherein a surface at alevel of the seat of the inlet valve and a facing conjugate surface at alevel of the bearing surface of the inlet valve have profiles having apoint of inflection, and wherein the line of contact moves in a vicinityof this point of inflection during pressure variations.
 15. An engineaccording to claim 14, wherein the at least one valve is the inletvalve, and wherein a transfer of heat occurs towards an outside surfaceof the inlet valve which is cooled cyclically by scavenging air.